WO2012160724A1

WO2012160724A1 - Internal combustion engine with variable compression ratio mechanism

Info

Publication number: WO2012160724A1
Application number: PCT/JP2011/075724
Authority: WO
Inventors: 坂柳　佳宏; 河崎　高志; 田中　宏幸; 敬野中井
Original assignee: トヨタ自動車株式会社
Priority date: 2011-05-23
Filing date: 2011-11-08
Publication date: 2012-11-29
Also published as: JP5569649B2; CN103547780B; JPWO2012160724A1; CN103547780A; US9644546B2; US20150128911A1

Abstract

This internal combustion engine is provided with a variable compression ratio mechanism capable of changing a mechanical compression ratio by changing the volume of the combustion chamber at the top dead center. The pressure and temperature of remaining combusted gas within the combustion chamber at the time when the exhaust valve is closed in the intake stroke is measured or estimated, the pressure and temperature of intake air supplied into the combustion chamber after the exhaust valve is closed in the intake stroke is measured or estimated, and based on the assumption that the pressure and temperature of the remaining combusted gas which saturates the volume of the combustion chamber at the time when the exhaust valve is closed in the air intake stroke become, when the intake air is supplied to the combustion chamber, equal to the pressure and temperature of the intake air, the volume of the remaining combusted gas after the change is calculated.

Description

Internal combustion engine having a variable compression ratio mechanism

The present invention relates to an internal combustion engine provided with a variable compression ratio mechanism.

2. Description of the Related Art An internal combustion engine having a variable compression ratio mechanism that varies a mechanical compression ratio by changing a combustion chamber volume at a top dead center is known. In such an internal combustion engine, the intake air amount may be controlled by making the closing timing of the intake valve variable in the compression stroke. In this case, it has been proposed to calculate the intake air amount based on the closing timing of the intake valve and the intake pressure (see Patent Document 1).

JP 2006-183604 A JP-A-2005-315161 JP 2005-233038 A JP 2004-111598 A JP2007-040212 JP 2010-265817 A JP 2009-092052 A

In an internal combustion engine equipped with the above-described variable compression ratio mechanism, the amount of change in the combustion chamber volume at the top dead center when the mechanical compression ratio is changed is assumed to be that only the amount of burnt gas remaining in the combustion chamber changes. The air volume is not affected. In other words, the intake air amount depends on the exhaust amount, and even if the mechanical compression ratio is changed, the exhaust amount does not change. Therefore, it has been considered that the intake air amount basically does not change. However, the residual burned gas volume shrinks due to a temperature drop when mixing with the intake air supplied into the combustion chamber, and expands due to a pressure drop, and thus affects the amount of intake air.

Therefore, an object of the present invention is to make it possible to estimate the intake air amount relatively accurately in an internal combustion engine having a variable compression ratio mechanism that changes the mechanical compression ratio by changing the volume of the combustion chamber at the top dead center. The volume of the remaining burned gas in the combustion chamber when the intake air is supplied to the combustion chamber is calculated.

An internal combustion engine comprising the variable compression ratio mechanism according to claim 1 of the present invention is an internal combustion engine comprising a variable compression ratio mechanism that varies the mechanical compression ratio by changing the volume of the combustion chamber at the top dead center. Measure or estimate the pressure and temperature of the residual burned gas in the combustion chamber when the exhaust valve is closed during the stroke, measure or estimate the pressure and temperature of the intake air supplied to the combustion chamber after the exhaust valve is closed during the intake stroke, When the pressure and temperature of the residual burned gas satisfying the combustion chamber volume when the exhaust valve is closed during the intake stroke are equal to the pressure and temperature of the intake when the intake air is supplied to the combustion chamber, The volume after change of the residual burned gas is calculated.

An internal combustion engine comprising the variable compression ratio mechanism according to claim 2 according to the present invention is an internal combustion engine comprising the variable compression ratio mechanism according to claim 1, and the intake air in the combustion chamber is based on the calculated volume of residual burned gas. The volume of fresh air is calculated and the volume of fresh air is calculated by multiplying the calculated volume of intake air by the fresh air ratio, assuming that burned gas is contained in the intake air supplied to the combustion chamber. And

According to the internal combustion engine having the variable compression ratio mechanism according to the first aspect of the present invention, the volume of the remaining burned gas in the combustion chamber when the exhaust valve is closed during the intake stroke is changed even when the intake air is supplied into the combustion chamber. Instead of occupying the combustion chamber volume when the exhaust valve is closed, the pressure and temperature of the residual burned gas become equal to the pressure and temperature of the intake air, and the volume of the remaining burned gas changes to occupy the combustion chamber volume. Therefore, the volume of the remaining burned gas after the change is calculated, and as a result, the volume of the combustion chamber occupied by the volume of the remaining burned gas is not supplied with intake air. The intake air amount can be estimated relatively accurately.

According to the internal combustion engine having the variable compression ratio mechanism according to claim 2 of the present invention, in the internal combustion engine having the variable compression ratio mechanism according to claim 1, the combustion chamber is based on the calculated residual burned gas volume. The intake air supplied to the combustion chamber is assumed to contain burnt gas, and the calculated intake volume is multiplied by the fresh air ratio to calculate the fresh air volume. It has become. Thereby, the amount of fresh air in the combustion chamber necessary for accurate calculation of the combustion air-fuel ratio can be estimated more accurately.

1 is an overall view of an internal combustion engine. It is a disassembled perspective view of a variable compression ratio mechanism. 1 is a schematic side sectional view of an internal combustion engine. It is a figure which shows a variable valve timing mechanism. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. It is a figure which shows the relationship between theoretical thermal efficiency and an expansion ratio. It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a flowchart for calculating the volume change of the residual burned gas in a combustion chamber.

FIG. 1 shows a side sectional view of an internal combustion engine equipped with a variable compression ratio mechanism according to the present invention. Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the

combustion chamber

5, and 7 is intake air. 8 is an intake port, 9 is an exhaust valve, and 10 is an exhaust port. The intake port 8 is connected to a surge tank 12 via an intake branch pipe 11, and a fuel injection valve 13 for injecting fuel into the corresponding intake port 8 is arranged in each intake branch pipe 11. The fuel injection valve 13 may be arranged in each combustion chamber 5 instead of being attached to each intake branch pipe 11.

The surge tank 12 is connected to an air cleaner 15 via an intake duct 14, and a throttle valve 17 driven by an actuator 16 and an intake air amount detector 18 using, for example, heat rays are arranged in the intake duct 14. On the other hand, the exhaust port 10 is connected to a catalyst device 20 containing, for example, a three-way catalyst via an exhaust manifold 19, and an air-fuel ratio sensor 21 is disposed in the exhaust manifold 19. When the combustion air-fuel ratio is the stoichiometric air-fuel ratio, it is preferable to incorporate a three-way catalyst in the catalyst device 20 as described above, but when the combustion air-fuel ratio is sometimes made leaner than the stoichiometric air-fuel ratio. it is preferably located downstream of the catalytic device 20 with a built-in three-way catalyst a different catalytic device containing a or _the NO _X storage reduction catalyst incorporating _the NO _X storage reduction catalyst to the catalytic converter 20.

On the other hand, in the embodiment shown in FIG. 1, the piston 4 is positioned at the compression top dead center by changing the relative position of the crankcase 1 and the cylinder block 2 in the cylinder axial direction at the connecting portion between the crankcase 1 and the cylinder block 2. There is provided a variable compression ratio mechanism A capable of changing the volume of the combustion chamber 5 at the time, and further an actual compression action start timing changing mechanism B capable of changing the actual start time of the compression action. In the embodiment shown in FIG. 1, the actual compression action start timing changing mechanism B is composed of a variable valve timing mechanism capable of controlling the closing timing of the intake valve 7.

As shown in FIG. 1, a relative position sensor 22 for detecting a relative positional relationship between the crankcase 1 and the cylinder block 2 is attached to the crankcase 1 and the cylinder block 2. Outputs an output signal indicating a change in the distance between the crankcase 1 and the cylinder block 2. The variable valve timing mechanism B is provided with a valve timing sensor 23 for generating an output signal indicating the closing timing of the intake valve 7, and an output signal indicating the throttle valve opening is provided to the actuator 16 for driving the throttle valve. A throttle opening sensor 24 is attached.

The electronic control unit 30 is composed of a digital computer, and is connected to each other by a bidirectional bus 31. A ROM (read only memory) 32, a RAM (random access memory) 33, a CPU (microprocessor) 34, an input port 35 and an output port 36. It comprises. Output signals of the intake air amount detector 18, the air-fuel ratio sensor 21, the relative position sensor 22, the valve timing sensor 23, and the throttle opening sensor 24 are input to the input port 35 via the corresponding AD converters 37. A load sensor 41 that generates an output voltage proportional to the depression amount L of the accelerator pedal 40 is connected to the accelerator pedal 40, and the output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. Is done. Further, a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 30 ° is connected to the input port 35. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injection valve 13, the throttle valve driving actuator 16, the variable compression ratio mechanism A, and the variable valve timing mechanism B through corresponding drive circuits 38.

2 shows an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 shows a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of protrusions 50 spaced from each other are formed below both side walls of the cylinder block 2, and cam insertion holes 51 each having a circular cross section are formed in each protrusion 50. Has been. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 1 so as to be fitted between the corresponding protrusions 50 spaced apart from each other. Cam insertion holes 53 each having a circular cross section are formed.

As shown in FIG. 2, a pair of

camshafts

54 and 55 are provided, and on each

camshaft

54 and 55, a circular cam 58 is rotatably inserted into each cam insertion hole 53. It is fixed. These circular cams 58 are coaxial with the rotational axes of the

camshafts

54 and 55. On the other hand, on both sides of each circular cam 58, as shown in FIG. 3, an eccentric shaft 57 eccentrically arranged with respect to the rotation axis of each

camshaft

54, 55 extends. 56 is mounted eccentrically and rotatable. As shown in FIG. 2, these circular cams 56 are arranged on both sides of each circular cam 58, and these circular cams 56 are rotatably inserted into the corresponding cam insertion holes 51. As shown in FIG. 2, a cam rotation angle sensor 25 that generates an output signal representing the rotation angle of the camshaft 55 is attached to the camshaft 55.

When the circular cams 58 fixed on the

camshafts

54 and 55 are rotated in opposite directions as indicated by arrows in FIG. 3A from the state shown in FIG. 3A, the eccentric shafts 57 are separated from each other. In order to move in the direction, the circular cam 56 rotates in the opposite direction to the circular cam 58 in the cam insertion hole 51, and as shown in FIG. 3B, the position of the eccentric shaft 57 is changed from the high position to the intermediate height position. It becomes. Next, when the circular cam 58 is further rotated in the direction indicated by the arrow, the eccentric shaft 57 is at the lowest position as shown in FIG.

3A, 3B, and 3C show the positional relationship among the center a of the circular cam 58, the center b of the eccentric shaft 57, and the center c of the circular cam 56 in each state. It is shown.

3A to 3C, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center a of the circular cam 58 and the center c of the circular cam 56. As the distance between the center a of 58 and the center c of the circular cam 56 increases, the cylinder block 2 moves away from the crankcase 1. That is, the variable compression ratio mechanism A changes the relative position between the crankcase 1 and the cylinder block 2 by a crank mechanism using a rotating cam. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the

camshafts

54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

As shown in FIG. 2, in order to rotate the

camshafts

54 and 55 in opposite directions, a pair of worms 61 and 62 having opposite spiral directions are attached to the rotation shaft of the drive motor 59, respectively.

Worm wheels

63 and 64 that mesh with the worms 61 and 62 are fixed to the ends of the

camshafts

54 and 55, respectively. In this embodiment, by driving the drive motor 59, the volume of the combustion chamber 5 when the piston 4 is located at the compression top dead center can be changed over a wide range.

On the other hand, FIG. 4 shows the variable valve timing mechanism B attached to the end of the camshaft 70 for driving the intake valve 7 in FIG. Referring to FIG. 4, the variable valve timing mechanism B includes a timing pulley 71 that is rotated in the direction of an arrow by a crankshaft of an engine via a timing belt, a cylindrical housing 72 that rotates together with the timing pulley 71, an intake valve A rotating shaft 73 that rotates together with the driving camshaft 70 and is rotatable relative to the cylindrical housing 72, and a plurality of partition walls 74 that extend from the inner peripheral surface of the cylindrical housing 72 to the outer peripheral surface of the rotating shaft 73. And a vane 75 extending from the outer peripheral surface of the rotating shaft 73 to the inner peripheral surface of the cylindrical housing 72 between the partition walls 74, and an advance hydraulic chamber 76 on each side of each vane 75. A retarding hydraulic chamber 77 is formed.

The hydraulic oil supply control to the

hydraulic chambers

76 and 77 is performed by the hydraulic oil supply control valve 78. The hydraulic oil supply control valve 78 includes

hydraulic ports

79 and 80 connected to the

hydraulic chambers

76 and 77, a hydraulic oil supply port 82 discharged from the hydraulic pump 81, a pair of

drain ports

83 and 84, And a spool valve 85 for controlling communication between the

ports

79, 80, 82, 83, and 84.

When the cam phase of the intake valve driving camshaft 70 should be advanced, the spool valve 85 is moved to the right in FIG. 4 and the hydraulic oil supplied from the supply port 82 is advanced via the hydraulic port 79. The hydraulic oil in the retard hydraulic chamber 77 is discharged from the drain port 84 while being supplied to the hydraulic chamber 76. At this time, the rotary shaft 73 is rotated relative to the cylindrical housing 72 in the direction of the arrow.

On the other hand, when the cam phase of the intake valve driving camshaft 70 should be retarded, the spool valve 85 is moved to the left in FIG. 4, and the hydraulic oil supplied from the supply port 82 causes the hydraulic port 80 to move. The hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83 while being supplied to the retard hydraulic chamber 77. At this time, the rotating shaft 73 is rotated relative to the cylindrical housing 72 in the direction opposite to the arrow.

If the spool valve 85 is returned to the neutral position shown in FIG. 4 while the rotation shaft 73 is rotated relative to the cylindrical housing 72, the relative rotation operation of the rotation shaft 73 is stopped, and the rotation shaft 73 is The relative rotation position at that time is held. Therefore, the variable valve timing mechanism B can advance and retard the cam phase of the intake valve driving camshaft 70 by a desired amount.

In FIG. 5, the solid line shows the time when the cam phase of the intake valve driving camshaft 70 is advanced the most by the variable valve timing mechanism B, and the broken line shows the cam phase of the intake valve driving camshaft 70 being the most advanced. It shows when it is retarded. Therefore, the valve opening period of the intake valve 7 can be arbitrarily set between the range indicated by the solid line and the range indicated by the broken line in FIG. 5, and therefore the closing timing of the intake valve 7 is also the range indicated by the arrow C in FIG. Any crank angle can be set.

The variable valve timing mechanism B shown in FIG. 1 and FIG. 4 shows an example. For example, the variable valve timing that can change only the closing timing of the intake valve while keeping the opening timing of the intake valve constant. Various types of variable valve timing mechanisms, such as mechanisms, can be used.

Next, the meaning of terms used in the present application will be described with reference to FIG. 6 (A), (B), and (C) show an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml for the sake of explanation. , (B), (C), the combustion chamber volume represents the volume of the combustion chamber when the piston is located at the compression top dead center.

FIG. 6A explains the mechanical compression ratio. The mechanical compression ratio is a value mechanically determined only from the stroke volume of the piston and the combustion chamber volume during the compression stroke, and this mechanical compression ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6A, this mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11.

FIG. 6B illustrates the actual compression ratio. This actual compression ratio is a value determined from the actual piston stroke volume and the combustion chamber volume from when the compression action is actually started until the piston reaches top dead center, and this actual compression ratio is (combustion chamber volume + actual (Stroke volume) / combustion chamber volume. That is, as shown in FIG. 6B, even if the piston starts to rise in the compression stroke, the compression operation is not performed while the intake valve is open, and the actual compression is performed from the time when the intake valve is closed. The action begins. Therefore, the actual compression ratio is expressed as described above using the actual stroke volume. In the example shown in FIG. 6B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

FIG. 6C explains the expansion ratio. The expansion ratio is a value determined from the stroke volume of the piston and the combustion chamber volume during the expansion stroke, and this expansion ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6C, this expansion ratio is (50 ml + 500 ml) / 50 ml = 11.

Next, the super expansion ratio cycle used in the present invention will be described with reference to FIGS. FIG. 7 shows the relationship between the theoretical thermal efficiency and the expansion ratio, and FIG. 8 shows a comparison between a normal cycle and an ultrahigh expansion ratio cycle that are selectively used according to the load in the present invention.

FIG. 8 (A) shows a normal cycle when the intake valve closes near the bottom dead center and the compression action by the piston is started from the vicinity of the intake bottom dead center. In the example shown in FIG. 8A as well, the combustion chamber volume is set to 50 ml, and the stroke volume of the piston is set to 500 ml, similarly to the example shown in FIGS. 6A, 6B, and 6C. As can be seen from FIG. 8A, in a normal cycle, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11, the actual compression ratio is almost 11, and the expansion ratio is also (50 ml + 500 ml) / 50 ml = 11. That is, in a normal internal combustion engine, the mechanical compression ratio, the actual compression ratio, and the expansion ratio are substantially equal.

The solid line in FIG. 7 shows the change in the theoretical thermal efficiency when the actual compression ratio and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, it is only necessary to increase the actual compression ratio. However, the actual compression ratio can only be increased to a maximum of about 12 due to the restriction of the occurrence of knocking at the time of engine high load operation, and thus the theoretical thermal efficiency cannot be sufficiently increased in a normal cycle.

On the other hand, under such circumstances, it is considered to increase the theoretical thermal efficiency while strictly dividing the mechanical compression ratio and the actual compression ratio. As a result, the theoretical thermal efficiency is governed by the expansion ratio, and the actual compression ratio is compared to the theoretical thermal efficiency. Was found to have little effect. That is, if the actual compression ratio is increased, the explosive force is increased, but a large amount of energy is required for compression. Thus, even if the actual compression ratio is increased, the theoretical thermal efficiency is hardly increased.

On the other hand, when the expansion ratio is increased, the period during which the pressing force acts on the piston during the expansion stroke becomes longer, and thus the period during which the piston applies the rotational force to the crankshaft becomes longer. Therefore, the larger the expansion ratio, the higher the theoretical thermal efficiency. The broken line ε = 10 in FIG. 7 indicates the theoretical thermal efficiency when the expansion ratio is increased with the actual compression ratio fixed at 10. Thus, the theoretical thermal efficiency when the expansion ratio is increased while the actual compression ratio ε is maintained at a low value, and the theoretical thermal efficiency when the actual compression ratio is increased with the expansion ratio as shown by the solid line in FIG. It can be seen that there is no significant difference from the amount of increase.

Thus, if the actual compression ratio is maintained at a low value, knocking does not occur. Therefore, if the expansion ratio is increased while the actual compression ratio is maintained at a low value, the theoretical thermal efficiency is reduced while preventing the occurrence of knocking. Can be greatly increased. FIG. 8B shows an example where the variable compression ratio mechanism A and variable valve timing mechanism B are used to increase the expansion ratio while maintaining the actual compression ratio at a low value.

Referring to FIG. 8 (B), in this example, the variable compression ratio mechanism A reduces the combustion chamber volume from 50 ml to 20 ml. On the other hand, the variable valve timing mechanism B delays the closing timing of the intake valve until the actual piston stroke volume is reduced from 500 ml to 200 ml. As a result, in this example, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11, and the expansion ratio is (20 ml + 500 ml) / 20 ml = 26. In the normal cycle shown in FIG. 8A, the actual compression ratio is almost 11 and the expansion ratio is 11, as described above. Compared with this case, only the expansion ratio is shown in FIG. 8B. It can be seen that it has been increased to 26. This is why it is called an ultra-high expansion ratio cycle.

Generally speaking, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency. Therefore, in order to improve the thermal efficiency during engine operation, that is, to improve fuel efficiency, it is necessary to improve the thermal efficiency when the engine load is low. Necessary. On the other hand, in the ultra-high expansion ratio cycle shown in FIG. 8B, since the actual piston stroke volume during the compression stroke is reduced, the amount of intake air that can be sucked into the combustion chamber 5 is reduced. The expansion ratio cycle can only be adopted when the engine load is relatively low. Therefore, in the present invention, when the engine load is relatively low, the super high expansion ratio cycle shown in FIG. 8B is used, and during the high engine load operation, the normal cycle shown in FIG. 8A is used.

Next, the overall operation control will be schematically described with reference to FIG. FIG. 9 shows changes in the intake air amount, the intake valve closing timing, the mechanical compression ratio, the expansion ratio, the actual compression ratio, and the opening degree of the throttle valve 17 according to the engine load at a certain engine speed. . 9 shows that the average air-fuel ratio in the combustion chamber 5 is an output signal of the air-fuel ratio sensor 21 so that unburned HC, CO and NO _x in the exhaust gas can be simultaneously reduced by the three-way catalyst in the catalyst device 20. This shows a case where feedback control is performed to the theoretical air-fuel ratio based on the above.

Now, as described above, the normal cycle shown in FIG. 8 (A) is executed during engine high load operation. Accordingly, as shown in FIG. 9, the expansion ratio is low because the mechanical compression ratio is lowered at this time, and the valve closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. ing. At this time, the amount of intake air is large, and at this time, the opening degree of the throttle valve 17 is kept fully open, so that the pumping loss is zero.

On the other hand, as shown by the solid line in FIG. 9, when the engine load becomes lower, the closing timing of the intake valve 7 is delayed in order to reduce the intake air amount. Further, at this time, as shown in FIG. 9, the mechanical compression ratio is increased as the engine load is lowered so that the actual compression ratio is kept substantially constant. Therefore, the expansion ratio is also increased as the engine load is lowered. At this time, the throttle valve 17 is kept fully open, and therefore the amount of intake air supplied into the combustion chamber 5 is controlled by changing the closing timing of the intake valve 7 regardless of the throttle valve 17. Has been.

Thus, when the engine load is reduced from the engine high load operation state, the mechanical compression ratio is increased as the intake air amount is decreased while the actual compression ratio is substantially constant. That is, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is decreased in proportion to the decrease in the intake air amount. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the intake air amount. At this time, in the example shown in FIG. 9, the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, so the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is proportional to the fuel amount. Will change.

When the engine load is further reduced, the mechanical compression ratio is further increased, and when the engine load is lowered to the medium load L1 slightly close to the low load, the mechanical compression ratio becomes a limit mechanical compression ratio (upper limit mechanical compression) that becomes the structural limit of the combustion chamber 5. Ratio). When the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio in a region where the load is lower than the engine load L1 when the mechanical compression ratio reaches the limit mechanical compression ratio. Accordingly, the mechanical compression ratio is maximized and the expansion ratio is maximized at the time of low engine load operation and low engine load operation, that is, at the engine low load operation side. In other words, the mechanical compression ratio is maximized so that the maximum expansion ratio is obtained on the engine low load operation side.

On the other hand, in the embodiment shown in FIG. 9, when the engine load decreases to L1, the closing timing of the intake valve 7 becomes the limit closing timing that can control the amount of intake air supplied into the combustion chamber 5. When the closing timing of the intake valve 7 reaches the limit closing timing, the closing timing of the intake valve 7 is reduced in a region where the load is lower than the engine load L1 when the closing timing of the intake valve 7 reaches the closing timing. It is held at the limit closing timing.

When the closing timing of the intake valve 7 is held at the limit closing timing, the intake air amount can no longer be controlled by the change in the closing timing of the intake valve 7. In the embodiment shown in FIG. 9, at this time, that is, in a region where the load is lower than the engine load L1 when the valve closing timing of the intake valve 7 reaches the limit valve closing timing, the intake valve 7 is supplied into the combustion chamber 5 by the throttle valve 17. The amount of intake air to be controlled is controlled, and the opening degree of the throttle valve 17 is made smaller as the engine load becomes lower.

On the other hand, the intake air amount can be controlled without depending on the throttle valve 17 by advancing the closing timing of the intake valve 7 as the engine load becomes lower as shown by the broken line in FIG. Accordingly, when expressing the case shown in FIG. 9 so as to include both the case indicated by the solid line and the case indicated by the broken line, in the embodiment according to the present invention, the valve closing timing of the intake valve 7 becomes smaller as the engine load becomes lower. It is moved in a direction away from the intake bottom dead center BDC until the limit valve closing timing L1 at which the intake air amount supplied into the combustion chamber can be controlled. Thus, the intake air amount can be controlled by changing the closing timing of the intake valve 7 as shown by the solid line in FIG. 9 or by changing it as shown by the broken line.

As described above, the expansion ratio is 26 in the ultra-high expansion ratio cycle shown in FIG. The higher the expansion ratio, the better. However, as can be seen from FIG. 7, a considerably high theoretical thermal efficiency can be obtained if it is 20 or more with respect to the practically usable lower limit actual compression ratio ε = 5. Therefore, in this embodiment, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more.

Incidentally, when determining the fuel injection amount, it is necessary to grasp the intake air amount (weight) supplied into the combustion chamber in order to realize a desired combustion air-fuel ratio. The intake air amount can be calculated based on the occupied volume of the intake air in the combustion chamber and the pressure and temperature of the intake air.

The occupied volume of the intake air in the combustion chamber is other than the occupied volume of the burned gas in the combustion chamber. In order to know the occupied volume of the intake air in the combustion chamber, the occupied volume of the burned gas in the combustion chamber may be calculated. . FIG. 10 is a flowchart for this purpose, and is implemented by the electronic control unit 30.

First, in step 101, it is determined whether or not it is time to determine the fuel injection amount. For example, when the fuel injection valve 13 is disposed in the intake port 8, the fuel injection is performed during the intake stroke. Further, when the fuel injection valve is disposed in the combustion chamber, fuel injection is possible from the initial stage of the intake stroke to the ignition timing of the compression stroke. However, in order to vaporize and mix the injected fuel, the fuel is injected during the intake stroke. It is preferable to end the injection. In any case, the fuel injection amount must be determined before the end of fuel injection.

When the determination in step 101 is negative, it is not necessary to calculate the occupied volume of the burned gas in the combustion chamber in order to determine the fuel injection amount, and the process ends without doing anything. However, if it is time to determine the fuel injection amount, the determination in step 101 is affirmed, and in step 102, the combustion chamber volume V0 when the exhaust valve is closed is set. The combustion chamber volume V0 when the exhaust valve is closed varies depending on not only the size and shape of the combustion chamber but also the current mechanical compression ratio and the current closing timing of the exhaust valve. As the mechanical compression ratio is reduced by the variable compression ratio mechanism A, the combustion chamber volume at the top dead center is increased, so that the combustion chamber volume V0 when the exhaust valve is closed is increased. Further, the more the exhaust valve closing timing is retarded, the larger the combustion chamber volume V0 when the exhaust valve is closed. The current mechanical compression ratio can be estimated based on the output of the relative position sensor 22.

During the intake stroke, even if the intake valve is opened and the piston descends from the intake top dead center (exhaust top dead center), the exhaust port 10 is in the exhaust port 10 while the exhaust valve is being opened due to the intake pressure of the intake port 8. Since the exhaust pressure is high, intake air is not supplied into the combustion chamber. Thus, the combustion chamber volume V0 when the exhaust valve is closed is filled with burned gas.

Next, in step 103, when the exhaust valve is closed, the temperature TEX and pressure PEX of the burned gas satisfying the combustion chamber volume V0 are measured by a temperature sensor and a pressure sensor (both not shown) arranged in the combustion chamber. The

Preferably, at the same time as step 103, in step 104, the temperature TIN and the pressure PIN of the intake air supplied to the combustion chamber are measured by, for example, a temperature sensor and a pressure sensor (both not shown) disposed in the surge tank 12. The

When the exhaust valve is closed during the intake stroke, the pressure PEX and the temperature TEX of the burned gas in the combustion chamber become equal to the intake pressure PIN and the temperature TIN when the intake air is supplied to the combustion chamber, and the volume of the burned gas changes. Then occupy the combustion chamber. In step 105, the volume V0 'of the burned gas that changes in this way is calculated by the following equation.

V0 '= V0 * TIN / TEX * PEX / PIN
Thus, if the occupied volume V0 ′ after the change in the volume of burned gas remaining in the combustion chamber is calculated immediately after closing the exhaust valve in the intake stroke, the intake air amount is calculated. can do. For example, when the valve closing timing of the intake valve is controlled before the intake bottom dead center as shown by a broken line in FIG. 9, the combustion chamber volume V1 ′ until the intake valve closes (based on the current mechanical compression ratio). The sum of the combustion chamber volume at the top dead center and the stroke volume of the piston from the top dead center to the intake valve closing) minus the occupied volume V0 'of burned gas (V1'-V0') is the intake occupation Thus, the intake air amount can be calculated based on the intake pressure PIN and the temperature TIN.

Further, when the valve closing timing of the intake valve is controlled after the intake bottom dead center as shown by the solid line in FIG. 9, for example, the combustion chamber volume V1 ″ (the current mechanical compression ratio) from the intake valve closing The volume (V1 ″ −V0 ′) obtained by subtracting the occupied volume V0 ′ of burned gas from the combustion chamber volume at the top dead center based on the above and the stroke volume of the piston from the intake valve closing to the top dead center) The amount of intake air can be calculated based on the intake pressure PIN and the temperature TIN.

Also, it can be considered that not only the intake air but also burned gas is discharged to the intake system between the intake bottom dead center and the intake valve closing. Subtract the combustion volume V0 'of the burned gas from the combustion chamber volume V1 at the intake bottom dead center (the sum of the combustion chamber volume at the top dead center and the piston stroke volume based on the current mechanical compression ratio). To calculate (V1-V0 ′). Next, the intake occupation volume (V1-V0 ′) at the intake bottom dead center is multiplied by the ratio V1 ″ / V1 between the combustion chamber volume V1 ″ from the intake valve closing and the combustion chamber volume V1 at the intake bottom dead center. By doing so, the occupied volume of the intake can be calculated, and the intake air amount can be calculated based on the intake pressure PIN and the temperature TIN.

In step 103 of the flowchart of FIG. 10, the burnt gas temperature TEX and pressure PEX in the combustion chamber when the intake valve is closed are measured by arranging a temperature sensor and a pressure sensor in the combustion chamber. Alternatively, it may be mapped for each engine operating state determined by the engine speed. Even if only the pressure sensor is arranged in the combustion chamber, the cylinder pressure P in the expansion stroke is monitored by the pressure sensor, and the product PV of the cylinder pressure P and the combustion chamber volume V becomes the maximum value PVM. By specifying the angle CA, it can be estimated that the burned gas temperature TEX becomes higher as the maximum value PVM is larger, and the later expansion work is reduced as the crank angle CA is retarded. Since TEX can be estimated to be high, the burned gas temperature TEX can be mapped to the maximum value PVM and the crank angle CA.

The intake air temperature TIN measured in step 104 of the flowchart of FIG. 10 may be the atmospheric temperature. Further, when the throttle valve is fully opened, the intake pressure PIN may be an atmospheric pressure. At the time of throttle valve opening control, the intake pressure PIN is mapped with respect to the throttle valve opening so that the intake valve PIN decreases as the throttle valve opening decreases (the absolute value of the negative pressure increases). It is also possible.

Further, when the intake valve opens during the exhaust stroke, the burned gas in the combustion chamber flows out not only to the exhaust port 10 but also to the intake port 8 from the intake valve open to the exhaust top dead center. Accordingly, strictly speaking, the intake air in the intake port 8 supplied from the exhaust valve closing into the combustion chamber contains burned gas.

Accordingly, the intake occupation volume (V1′−V0 ′), (V1 ″ −V0 ′), or (V1−V0 ′) · V ″ / V1 calculated as described above, and from the intake port 8 into the combustion chamber. By multiplying the intake air fresh air ratio R, the fresh air occupation volume for estimating the amount of fresh air necessary for accurate calculation of the combustion air-fuel ratio can be calculated.

Here, the fresh air ratio R is a ratio fv / gv of the fresh air volume fv to the unit volume gv of the gas sucked into the combustion chamber from the intake port 9, and the unit volume gv is the fresh air included in the unit volume gv. It is the sum of the volume fv and the burned gas volume ev.

As the valve opening timing in the exhaust stroke of the intake valve is advanced, the amount of burned gas flowing out to the intake port 8 increases, and the burned gas with respect to the unit volume gv of gas sucked into the combustion chamber from the intake port 9 Since the volume ev increases, the fresh air ratio R decreases. In addition, the higher the engine load and the higher the combustion pressure, the higher the burnt gas pressure in the cylinder when the intake valve is opened, so the amount of burnt gas flowing out to the intake port 8 increases and the fresh air ratio R Becomes smaller. Thus, the fresh air ratio R can be mapped based on the engine operating state (engine load and engine speed) and the opening timing of the intake valve.

1 Crankcase 2 Cylinder block A Variable compression ratio mechanism B Variable valve timing mechanism

Claims

An internal combustion engine having a variable compression ratio mechanism that varies the compression chamber ratio by changing the combustion chamber volume at the top dead center, and the pressure and temperature of the remaining burned gas in the combustion chamber when the exhaust valve is closed during the intake stroke The residual burned gas satisfying the combustion chamber volume when the exhaust valve is closed during the intake stroke by measuring or estimating the pressure and temperature of the intake air supplied to the combustion chamber after the exhaust valve is closed during the intake stroke The volume after the change of the residual burned gas is calculated on the assumption that the pressure and the temperature of the exhaust gas are equal to the pressure and the temperature of the intake air when the intake air is supplied to the combustion chamber. An internal combustion engine having a ratio mechanism.
Calculate the intake air volume in the combustion chamber based on the calculated residual burned gas volume, and assume that the intake air supplied to the combustion chamber contains burnt gas, and the fresh air ratio in the calculated intake air volume The internal combustion engine provided with the variable compression ratio mechanism according to claim 1, wherein the volume of fresh air is calculated by multiplying.